JP3463576B2 - Internal combustion engine - Google Patents

Description

Detailed Description of the Invention

[0001]

TECHNICAL FIELD The present invention relates to an internal combustion engine.

[0002]

2. Description of the Related Art Conventionally, in an internal combustion engine, for example, a diesel engine, an engine exhaust passage and an engine intake passage are connected by an exhaust gas recirculation (hereinafter referred to as EGR) passage in order to suppress the generation of NO _x. Exhaust gas, that is, EGR gas, is recirculated into the engine intake passage via the EGR passage. In this case, the EGR gas has a relatively high specific heat and therefore can absorb a large amount of heat.
The combustion temperature in the combustion chamber decreases as the GR gas amount increases, that is, as the EGR rate (EGR gas amount / (EGR gas amount + intake air amount)) increases. When the combustion temperature decreases, the amount of NO _x generated decreases. Therefore, the higher the EGR rate, the lower the amount of NO _x generated.

As described above, it has been known that the amount of NO _x generated can be reduced by increasing the EGR rate. However, when the EGR rate is increased, when the EGR rate exceeds a certain limit, the amount of soot generated, that is, the smoke starts to increase rapidly. In this regard, it has been conventionally thought that if the EGR rate is further increased, the smoke will increase infinitely, and therefore the smoke will start to increase rapidly.
The GR rate is considered to be the maximum allowable limit for the EGR rate.

Therefore, conventionally, the EGR rate is set within a range that does not exceed the maximum allowable limit. The maximum allowable limit of this EGR rate is approximately 30 to 50 percent, though it varies considerably depending on the engine type and fuel.
Therefore, in the conventional diesel engine, the maximum EGR rate is 3
It is suppressed from 0% to 50%.

As described above, in the past, it was considered that the maximum allowable limit exists for the EGR rate.
The R rate is NO within the range that does not exceed this maximum allowable limit.
_It was stipulated that the amount of _x and smoke generated should be as small as possible. However, in this way EGR
Even if the rate is set so that the amount of NO _x and smoke produced is as small as possible, there is a limit to the reduction in the amount of NO _x and smoke produced, and in reality, a considerable amount of N 2 is still left.
The O _x, and smoke is generated at present.

However, if the EGR rate is made larger than the maximum permissible limit in the process of studying the combustion of a diesel engine, the smoke increases sharply as described above, but there is a peak in the amount of smoke produced, and the peak is exceeded. EGR
When the rate is further increased, the smoke starts to decrease sharply this time. When the EGR rate is 70% or more during idling operation, and when the EGR gas is strongly cooled, the EGR rate is almost 55% or more. It was found that it was almost zero, that is, soot was hardly generated. At this time, N
It has also been found that the amount of O _x generated is extremely small.
After that, the reason why soot was not generated was examined based on this finding, and as a result, soot and NO
_This led to the construction of a new combustion system capable of simultaneously reducing _x . This new combustion system will be explained in detail later, but in short, it is basically based on stopping the growth of hydrocarbons in the middle of the process until the hydrocarbons grow into soot.

[0007] That is, as a result of repeated experimental research, it was found that when the temperature of the fuel and the gas around it during combustion in the combustion chamber were below a certain temperature, the growth of hydrocarbons stopped in the middle of the process before reaching soot. However, if the temperature of the fuel and the gas around it rises above a certain temperature, the hydrocarbons will suddenly grow to soot. In this case, the temperature of the fuel and its surrounding gas is greatly affected by the endothermic action of the gas around the fuel when the fuel burns, and the endothermic amount of the gas around the fuel is adjusted according to the amount of heat generated during fuel combustion. Thus, the temperature of the fuel and the gas around it can be controlled.

Therefore, if the temperature of the fuel and the gas around it during combustion in the combustion chamber is suppressed below the temperature at which the growth of hydrocarbons stops halfway, soot will not be generated and the fuel and the surroundings during combustion in the combustion chamber will be eliminated. It is possible to control the gas temperature in the range below the temperature at which the growth of hydrocarbons stops halfway by adjusting the heat absorption amount of the gas around the fuel. On the other hand, hydrocarbons whose growth has stopped before reaching soot can be easily purified by post-treatment using an oxidation catalyst or the like. This is the basic idea of the new combustion system. The applicant has already applied for an internal combustion engine that employs this new combustion system (Japanese Patent Application No. 9-305850).

[0009]

By the way, under this new combustion, the EGR gas amount has a great influence on the combustion.
Thus, there is an optimum EGR rate that allows good combustion for the engine operating conditions. Therefore, in order to ensure good combustion under this new combustion, it is necessary to control the EGR rate to an optimum EGR rate according to the operating state of the engine.

By the way, since the EGR rate is expressed by the EGR gas amount / (EGR gas amount + intake air amount) as described above, EG
If the R gas amount and the intake air amount are known, the EGR rate can be obtained. If the EGR rate can be obtained, the EGR rate can be controlled by controlling the EGR gas amount or the intake air amount, or both the EGR gas amount and the intake air amount. Can be controlled to an optimum EGR rate according to the operating state of the engine.

However, in this case, as a practical matter, it is difficult to detect the amount of EGR gas that is recirculated in the intake passage, and therefore, by some other method, EG
It is necessary to control the R rate to an optimum EGR rate according to the operating state of the engine.

[0012]

In [SUMMARY OF THE Therefore the first invention, it reaches a peak generation amount of As you increase the recirculated exhaust gas amount supplied to the combustion chamber soot increases gradually, the combustion
Further increase the amount of recirculated exhaust gas supplied to the room
And fuel around it during combustion in the chamber and combustion chamber
The temperature is lower than the soot formation temperature and soot is almost generated.
In an internal combustion engine that eliminates combustion by increasing the recirculated exhaust gas where the amount of production of soot is supplied to the combustion chamber than the recirculation amount of exhaust gas reaches a peak supplies recirculated exhaust gas into the engine intake passage When burning indoors
The temperature of the fuel and the gas around it is the temperature at which soot is produced.
The intake air amount control means for suppressing the intake air amount supplied to the combustion chamber to a predetermined intake air amount according to the operating state of the engine, and the pressure in the intake passage. And a recirculation exhaust gas amount control unit for controlling the recirculation exhaust gas amount so that the exhaust gas recirculation ratio becomes the target exhaust gas recirculation ratio based on the pressure in the intake passage. ing.

That is, when the intake air amount or EGR gas amount decreases, the pressure in the intake passage decreases, whereas when the intake air amount or EGR gas amount increases, the intake passage pressure increases. That is, the pressure in the intake passage is determined by the sum of the intake air amount and the EGR gas amount. In the first aspect, the intake air amount control means controls the intake air amount to a predetermined intake air amount. Therefore, at this time, the EGR gas amount from the pressure in the intake passage is equal to the predetermined EGR gas amount. Whether or not there is, that is, whether or not the EGR rate is the target EGR rate can be determined.

In this case, when the pressure in the intake passage is lower than the pressure in the intake passage when the EGR rate is the target EGR rate, the EGR gas amount is insufficient, and in this case. If the EGR gas amount is increased, the EGR rate becomes the target EGR rate. On the other hand, when the pressure in the intake passage is higher than the pressure in the intake passage when the EGR rate is the target EGR rate, the EGR gas amount is excessive. In this case, the EGR gas amount is If is decreased, the EGR rate becomes the target EGR rate. That is, the EGR rate is controlled by controlling the EGR amount based on the pressure in the intake passage.
The R rate can be controlled.

In the second invention, in the first invention,
The exhaust gas recirculation rate is set to a target exhaust gas recirculation rate. The target pressure in the intake passage is stored in advance according to the engine operating condition necessary for setting the target exhaust gas recirculation rate. The amount of recirculated exhaust gas is controlled so that the internal pressure becomes the target pressure. 2 in the third invention
In the second invention, the recirculation exhaust gas amount control means comprises an exhaust gas recirculation control valve arranged in the exhaust gas recirculation passage connecting the engine exhaust passage and the engine intake passage, and the recirculation exhaust gas amount control The means controls the opening degree of the exhaust gas recirculation control valve so that the pressure in the intake passage becomes the target pressure.

In the fourth invention, in the first invention,
The intake air amount detecting means for detecting the intake air amount, and the air-fuel ratio control means for controlling the air-fuel ratio to the target air-fuel ratio according to the operating state of the engine based on the detected intake air amount are provided. . A fifth aspect of the invention is the fuel cell system according to the fourth aspect of the invention, further including a calculating unit that calculates the fuel injection amount, and the air-fuel ratio control unit sets the ratio between the detected intake air amount and the calculated fuel injection amount to the target air-fuel ratio. In this way, the intake air amount is corrected.

In the sixth invention, in the fourth invention,
The air-fuel ratio detection means for detecting the air-fuel ratio is provided, and the air-fuel ratio control means corrects the intake air amount so that the detected air-fuel ratio becomes the target air-fuel ratio. In the seventh invention, in the fourth invention, the air-fuel ratio control means calculates the fuel injection amount necessary for making the air-fuel ratio the target air-fuel ratio based on the detected intake air amount.

In the eighth invention, in the first invention,
The intake air amount control means includes a throttle valve arranged in the engine intake passage, and the throttle valve is configured so that the intake air amount supplied into the combustion chamber becomes a predetermined intake air amount according to the operating state of the engine. The opening degree of is controlled. According to a ninth aspect of the present invention, in the first aspect of the present invention, a storage unit that stores the injection timing is provided, and the injection timing is controlled based on the stored injection timing.

In the tenth aspect of the present invention, in the ninth aspect, a combustion pressure detecting means for detecting the combustion pressure in the combustion chamber is provided, and the injection timing is controlled based on this combustion pressure. The eleventh aspect of the present invention is the first aspect of the present invention, which is provided with a rotational speed control means for controlling the engine rotational speed to a target idling rotational speed during engine idling operation.

According to the twelfth aspect of the invention, in the eleventh aspect of the invention, the rotation speed control means controls the fuel injection amount to maintain the engine rotation speed at the target idling rotation speed. According to a thirteenth aspect of the invention, in the eleventh aspect of the invention, the rotation speed control means controls the intake air amount to maintain the engine rotation speed at the target idling rotation speed.

According to a fourteenth invention, in the first invention, the intake air amount control means includes a throttle valve arranged in the engine intake passage, and the intake passage inner wall surface portion around the throttle valve is expanded outward. As the throttle valve opens from the fully closed position, the flow passage area between the valve body peripheral edge of the throttle valve and the intake passage inner wall surface portion is increased as the throttle valve opens from the fully closed position. The shape is such that it gradually increases.

According to the fifteenth invention, in the first invention, the exhaust gas recirculation rate is about 55% or more. According to the sixteenth invention, in the first invention, a catalyst having an oxidation function is arranged in the engine exhaust passage. 17
In the 16th aspect of the invention, the catalyst comprises at least one of an oxidation catalyst, a three-way catalyst and an NO _x absorbent.

According to the eighteenth invention, in the first invention, the recirculation exhaust gas amount supplied to the combustion chamber is larger than the recirculation exhaust gas amount at which the soot generation amount reaches a peak, and soot is hardly generated. A switching means is provided for selectively switching between combustion and second combustion in which the amount of recirculated exhaust gas supplied into the combustion chamber is smaller than the amount of recirculated gas at which the amount of soot generated peaks.

According to a nineteenth invention, in the eighteenth invention, the engine operating region is divided into a first operating region on the low load side and a second operating region on the high load side, and the first operating region is divided into the first operating region and the first operating region. Is performed, and the second combustion is performed in the second operation region.

[0025]

FIG. 1 shows the case where the present invention is applied to a four-stroke compression ignition type internal combustion engine. Referring to FIG. 1, 1 is an engine body, 2 is a cylinder block, 3 is a cylinder head, 4 is a piston, 5 is a combustion chamber, 6 is an electrically controlled fuel injection valve, 7 is an intake valve, 8 is an intake port, and 9 is an intake port. Indicates an exhaust valve, and 10 indicates an exhaust port, respectively. The intake port 8 is connected to a surge tank 12 via a corresponding intake branch pipe 11, and the surge tank 12 is connected via an intake duct 13 and an intercooler 14 to a supercharger, for example, an outlet portion of a compressor 16 of an exhaust turbocharger 15. Be connected. The inlet of the compressor 16 is connected to an air cleaner 18 via an air suction pipe 17, and a throttle valve 20 driven by a step motor 19 is arranged in the air suction pipe 17. In addition, the air suction pipe 17 upstream of the throttle valve 20.
A mass flow rate detector 21 for detecting the mass flow rate of the intake air is arranged therein.

On the other hand, the exhaust port 10 is connected to the exhaust manifold 2
Exhaust turbine 2 of exhaust turbocharger 15 via 2
3 and the outlet of the exhaust turbine 23 is connected via an exhaust pipe 24 to a catalytic converter 26 containing a catalyst 25 having an oxidizing function. Exhaust manifold 22
An air-fuel ratio sensor 27 is arranged inside. The exhaust pipe 28 connected to the outlet of the catalytic converter 26 and the throttle valve 2
The downstream air suction pipe 17 is connected to each other via an EGR passage 29, and the step motor 3 is provided in the EGR passage 29.
An EGR control valve 31 driven by 0 is arranged. Further, in the EGR passage 29, E flowing in the EGR passage 29
An intercooler 32 for cooling the GR gas is arranged. In the embodiment shown in FIG. 1, the engine cooling water is guided into the intercooler 32, and the engine cooling water cools the EGR gas.

On the other hand, fuel injectors 6 are connected through a fuel supply pipe 33 the fuel reservoir, a so-called common rail 34. The common rail is to Le 34 is supplied with fuel from a variable discharge fuel pump 35 of the electrically controlled, common rail
The fuel supplied into the fuel cell 34 is supplied to the fuel injection valve 6 via each fuel supply pipe 33. The common rail 34 is mounted a fuel pressure sensor 36 for detecting the fuel pressure in Como <br/> Nre Le 34, the target fuel fuel pressure common rail 34 based on the output signal of the fuel pressure sensor 36 The discharge amount of the fuel pump 35 is controlled so that the pressure becomes a pressure.

The electronic control unit 40 is composed of a digital computer, and has a ROM (Read Only Memory) 42, a RAM (Random Access Memory) 43, a CPU (Microprocessor) 44, an input port 45, and an input port 45 which are connected to each other by a bidirectional bus 41. The output port 46 is provided. The output signal of the mass flow rate detector 21 is input to the input port 45 via the corresponding AD converter 47, and the output signals of the air-fuel ratio sensor 27 and the fuel pressure sensor 36 are also input ports via the corresponding AD converter 47. 45 is input. The combustion chamber 5 is arranged a combustion pressure sensor 3 7 for detecting the pressure in the combustion chamber 5, the output signal of the combustion pressure sensor 3 7 is connected to the input terminal I of a peak hold circuit 49. The output terminal O of the peak hold circuit 49 is input to the input port 45 via the corresponding AD converter 47. Further, the throttle valve 20 downstream of the air suction pipe 17 is mounted a pressure sensor 3 8 for detecting the absolute pressure in the air suction pipe 17, AD converter 47 the output signal of the pressure sensor 3 8 corresponding Is input to the input port 45 via.

A load sensor 51 for generating an output voltage proportional to the depression amount L of the accelerator pedal 50 is connected to the accelerator pedal 50, and the output voltage of the load sensor 51 is input to the input port via a corresponding AD converter 47. 45 is input. A crank angle sensor 52 that generates an output pulse each time the crankshaft rotates, for example, 30 ° is connected to the input port 45. Further, the output pulse of the vehicle speed sensor 53 indicating the vehicle speed is input to the input port 45. On the other hand, the output port 46 is connected to the reset input terminal R of the fuel injection valve 6, the throttle valve control step motor 19, the EGR control valve control step motor 30, the fuel pump 35 and the peak hold circuit 49 via the corresponding drive circuit 48. Connected.

FIG. 2 shows an enlarged view around the throttle valve 20. Referring to FIG. 2, the inner wall portion 17a of the air suction pipe 17 around the throttle valve 20, i.e., the intake passage 3 9
The inner wall portion 17a of which is caused to bulge outward, the inner wall portion of the intake passage 3 9 in the embodiment shown in FIG. 2 1
7a has an arcuate cross-sectional shape. When the throttle valve 20 is opened from the fully closed position as compared with the case in this way the cross-sectional shape of the inner wall portion 17a of the intake passage 3 9 air suction pipe 17 when the arc shape throughout having a cylindrical shape The flow passage area between the valve body peripheral edge of the throttle valve 20 and the inner wall surface portion 17a gradually increases. In other words, the change in the flow passage area with respect to the rotation angle of the throttle valve 20 is smaller than in the case where the air suction pipe 17 has a cylindrical shape as a whole. As a result, it becomes possible to finely control the intake air amount by the throttle valve 20.

[0031] shows another embodiment of the shape of the inner wall portion 17a of the intake passage 3 9 3. Intake passage 3 9 In this example
The cross-sectional shape of the inner wall surface portion 17a of V is V-shaped. Also in this embodiment, the change in the flow passage area with respect to the rotation angle of the throttle valve 20 is smaller than that in the case where the air suction pipe 17 has a cylindrical shape as a whole.
It is possible to finely control the intake air amount by zero.

FIG. 4 shows the throttle valve 2 during engine low load operation.
The change in output torque when the air-fuel ratio A / F (horizontal axis in FIG. 4) is changed by changing the opening degree of 0 and the EGR rate, and the change in the amount of smoke, HC, CO, and NO _x emissions. It represents the experimental example shown. As can be seen from FIG. 4, in this experimental example, the EGR rate increases as the air-fuel ratio A / F decreases, and when the air-fuel ratio is equal to or less than the theoretical air-fuel ratio (≈14.6), the EGR rate is 65% or more.

As shown in FIG. 4, when the air-fuel ratio A / F is reduced by increasing the EGR rate, the EGR rate becomes around 40%, and when the air-fuel ratio A / F reaches about 30, smoke is generated. The amount of generation begins to increase. Next, when the EGR rate is further increased and the air-fuel ratio A / F is made smaller, the amount of smoke generated sharply increases and reaches a peak. Next, when the EGR rate is further increased and the air-fuel ratio A / F is reduced, the smoke sharply decreases this time, the EGR rate is increased to 65% or more, and the smoke becomes almost zero when the air-fuel ratio A / F is around 15.0. . That is, soot is hardly generated. At this time, the output torque of the engine slightly decreases, and N
The amount of O _x generated is considerably low. On the other hand, at this time, HC,
The amount of CO generated starts to increase.

FIG. 5 (A) shows the change in combustion pressure in the combustion chamber 5 when the air-fuel ratio A / F is around 21 and the amount of smoke generated is the largest, and FIG. 5 (B) shows the air-fuel ratio A / F. It shows a change in the combustion pressure in the combustion chamber 5 when F is around 18 and the amount of smoke generated is almost zero. As can be seen by comparing FIG. 5 (A) and FIG. 5 (B), in the case shown in FIG. 5 (B) where the amount of smoke generated is almost zero, the amount of smoke generated is large.
It can be seen that the combustion pressure is lower than in the case shown in (A).

From the experimental results shown in FIGS. 4 and 5, the following can be said. That is, first of all, the air-fuel ratio A / F is 1
When the amount of smoke generated is 5.0 or less and the amount of smoke is almost zero, FIG.
As shown in (3), the amount of NO _x generated is considerably reduced. N
The decrease in the amount of generated O _x means that the combustion temperature in the combustion chamber 5 is decreased, and therefore, when the soot is hardly generated, the combustion temperature in the combustion chamber 5 is decreased. I can say. The same can be said from FIG. That is, the combustion pressure is low in the state shown in FIG. 5 (B) in which almost no soot is generated.
The combustion temperature inside is low.

Second, when the amount of smoke produced, that is, the amount of soot produced, becomes almost zero, as shown in FIG.
Emissions will increase. This means that hydrocarbons are discharged without growing to soot. That is, the linear hydrocarbons and aromatic hydrocarbons contained in the fuel as shown in FIG. 6 are thermally decomposed to form soot precursors when the temperature is raised in a state of oxygen deficiency, and then mainly soot is formed. Soot consisting of a solid with carbon atoms gathered is produced. In this case, the actual soot formation process is complicated, and it is not clear what form the soot precursor takes, but in any case, the hydrocarbon as shown in FIG. After that, it will grow to soot. Therefore, as described above, when the soot generation amount becomes almost zero, the HC and CO emission amounts increase as shown in FIG. 4, but at this time, the HC is a soot precursor or a hydrocarbon in the state before it. .

When these considerations based on the experimental results shown in FIGS. 4 and 5 are summarized, the soot generation amount becomes almost zero when the combustion temperature in the combustion chamber 5 is low, and at this time, the soot precursor or the soot precursor The hydrocarbons in this state are discharged from the combustion chamber 5. As a result of further detailed experimental research on this, when the temperature of the fuel and the gas around it in the combustion chamber 5 is below a certain temperature, the soot growth process stops halfway, that is, the soot is generated. It was found that soot was not generated at all and soot was generated when the temperature of the fuel and its surroundings in the combustion chamber 5 reached a certain temperature or higher.

By the way, the temperature of the fuel and its surroundings when the hydrocarbon production process stops in the state of the soot precursor, that is, the above-mentioned certain temperature depends on various factors such as the type of fuel, the air-fuel ratio and the compression ratio. It cannot be said how many times it changes, but this certain temperature has a deep relationship with the amount of NO _x produced, and therefore this certain temperature is defined to some extent from the amount of NO _x produced. be able to. That is, as the EGR rate increases, the temperature of the fuel during combustion and the gas around it decreases, and the amount of NO _x generated decreases. At this time, soot is hardly generated when the amount of NO _x generated is about 10 p.pm or less. Therefore, the above certain temperature is NO
It is almost the same as the temperature when the amount of _x generation is around 10 p.pm or less.

Once soot is produced, this soot cannot be purified by a post-treatment using a catalyst having an oxidizing function. On the other hand, the soot precursor or the hydrocarbon in the state before it can be easily purified by a post-treatment using a catalyst having an oxidizing function. Considering the post-treatment with a catalyst having an oxidizing function as described above, it is extremely difficult to determine whether the hydrocarbon is discharged from the combustion chamber 5 in the state of the soot precursor or in the state before it, or is discharged from the combustion chamber 5 in the form of soot. There is a big difference. The new combustion system employed in the present invention allows hydrocarbons to be discharged from the combustion chamber 5 in the form of soot precursors or pre-presence conditions without producing soot in the combustion chamber 5 The core is to oxidize with a catalyst having an oxidizing function.

Now, in order to stop the growth of hydrocarbons before the soot is generated, the temperature of the fuel and the gas around it in the combustion chamber 5 during combustion is set to a temperature lower than the temperature at which the soot is generated. It needs to be suppressed. In this case, it has been found that, in order to suppress the temperature of the fuel and the gas around it, the endothermic action of the gas around the fuel when the fuel burns has an extremely large effect.

That is, when only air exists around the fuel, the evaporated fuel immediately reacts with oxygen in the air and burns. In this case, the temperature of the air separated from the fuel does not rise so much, and only the temperature around the fuel locally becomes extremely high. That is, at this time, the air separated from the fuel hardly absorbs the combustion heat of the fuel. In this case, since the combustion temperature locally becomes extremely high, the unburned hydrocarbons that have received this heat of combustion generate soot.

On the other hand, the situation is slightly different when the fuel is present in a mixed gas of a large amount of inert gas and a small amount of air.
In this case, the evaporated fuel diffuses into the surroundings, reacts with oxygen mixed in the inert gas, and burns. In this case, the combustion heat is absorbed by the surrounding inert gas, so that the combustion temperature does not rise so much. That is, the combustion temperature can be kept low. That is, the presence of the inert gas plays an important role in suppressing the combustion temperature, and the combustion temperature can be suppressed low by the endothermic action of the inert gas.

In this case, in order to suppress the temperature of the fuel and the gas around it to a temperature lower than the temperature at which soot is produced, an amount of inert gas sufficient to absorb the amount of heat required to do so is required. . Therefore, if the fuel amount increases, the required amount of inert gas also increases accordingly. In this case, the larger the specific heat of the inert gas, the stronger the endothermic action. Therefore, the inert gas is preferably a gas having a large specific heat. In this respect, since CO ₂ and EGR gas have a relatively large specific heat, it can be said that it is preferable to use EGR gas as the inert gas.

FIG. 7 shows the relationship between the EGR rate and smoke when EGR gas is used as the inert gas and the cooling degree of the EGR gas is changed. That is, the curve A in FIG. 7 strongly cools the EGR gas to bring the EGR gas temperature to about 9
The curve B shows the case where the EGR gas is cooled by a small cooling device, and the curve C shows the case where the temperature is maintained at 0 ° C.
Indicates the case where the EGR gas is not forcibly cooled.

As shown by the curve A in FIG. 7, when the EGR gas is strongly cooled, the soot generation amount peaks when the EGR rate is slightly lower than 50%, and in this case, the EGR rate is almost 55. Almost no soot is generated if the percentage is exceeded. On the other hand, as shown by the curve B in FIG. 7, when the EGR gas is slightly cooled, the soot generation amount reaches a peak when the EGR rate is slightly higher than 50%, and in this case, the EGR rate is approximately 65% or more. If so, soot is hardly generated.

Further, as shown by the curve C in FIG. 7, EG
When the R gas is not forcibly cooled, the EGR rate is 5
The soot generation amount peaks near 5%, and in this case, if the EGR rate is set to approximately 70% or more, soot is hardly generated. Note that FIG. 7 shows the amount of smoke generated when the engine load is relatively high, and the EGR rate at which the amount of soot generated peaks when the engine load decreases and the EGR rate at which soot hardly occurs is slightly decreased. The lower limit of is also slightly lowered. Thus, the lower limit of the EGR rate at which soot is hardly generated changes depending on the cooling degree of EGR gas and the engine load.

FIG. 8 shows a mixture of EGR gas and air required to bring the temperature of the fuel and its surrounding gas at the time of combustion to a temperature lower than the temperature at which soot is generated when EGR gas is used as the inert gas. The amount of gas, the ratio of air in this mixed gas amount, and the ratio of EGR gas in this mixed gas are shown. In addition, in FIG. 8, the vertical axis represents the total intake gas amount sucked into the combustion chamber 5, and the chain line Y represents the total intake gas amount that can be sucked into the combustion chamber 5 when supercharging is not performed. ing. The horizontal axis shows the required load.

Referring to FIG. 8, the proportion of air, that is, the amount of air in the mixed gas indicates the amount of air required to completely burn the injected fuel. That is, in the case shown in FIG. 8, the ratio between the air amount and the injected fuel amount is the stoichiometric air-fuel ratio. On the other hand, in FIG. 8, the ratio of the EGR gas, that is, the amount of the EGR gas in the mixed gas is set so that the temperature of the fuel and its surrounding gas becomes lower than the temperature at which soot is formed when the injected fuel is burned. The minimum required EGR gas amount is shown. This EGR gas amount is approximately 55% or more when expressed by the EGR rate, and is 7 in the embodiment shown in FIG.
It is 0% or more. That is, the total intake gas amount sucked into the combustion chamber 5 is shown by the solid line X in FIG. 8, and the ratio of the air amount to the EGR gas amount in this total intake gas amount X is shown in FIG.
When the ratio is as shown in (1), the temperature of the fuel and the gas around it becomes lower than the temperature at which soot is generated, and thus soot is not generated at all. Further, the amount of NO _x generated at this time is around 10 p.pm or less, so N
The amount of O _x generated is extremely small.

When the fuel injection amount increases, the amount of heat generated when the fuel burns also increases. Therefore, in order to maintain the temperature of the fuel and the gas around it at a temperature lower than the temperature at which soot is generated, heat generated by the EGR gas is used. The amount of absorption must be increased. Therefore, as shown in FIG. 8, the EGR gas amount must be increased as the injected fuel amount is increased.
That is, the EGR gas amount needs to increase as the required load increases.

By the way, when the supercharging is not performed, the upper limit of the total intake gas amount X sucked into the combustion chamber 5 is Y. Therefore, in FIG. 8, the required load is larger than the required load L _0. The air-fuel ratio cannot be maintained at the stoichiometric air-fuel ratio unless the EGR gas ratio is decreased with increasing. In other words, when supercharging is not performed and the air-fuel ratio is kept at the stoichiometric air-fuel ratio in the region where the required load is larger than L ₀ , the EGR rate decreases as the required load increases, and Therefore, in the region where the required load is larger than L _{0, the} temperature of the fuel and the gas around it cannot be maintained at a temperature lower than the temperature at which soot is generated. However, as shown in FIG. 1, when EGR gas is recirculated to the inlet side of the supercharger, that is, the air suction pipe 17 of the exhaust turbocharger 15 via the EGR passage 29, the required load is L _0.
In the larger region, the EGR rate can be maintained above 55 percent, for example 70 percent, thus maintaining the fuel and surrounding gas temperatures below the temperature at which soot is produced. That is, if the EGR gas is recirculated so that the EGR rate in the air suction pipe 17 becomes, for example, 70%, the EGR rate of the intake gas boosted by the compressor 16 of the exhaust turbocharger 15 also becomes 70%. The temperature of the fuel and the gas around the fuel can be maintained at a temperature lower than the temperature at which the soot is generated, up to the limit where the pressure can be boosted by the compressor 16. Therefore, the operating range of the engine capable of producing the low temperature combustion can be expanded. When the EGR rate is set to 55% or more in the region where the required load is larger than L ₀ , the EGR control valve 31 is fully opened and the throttle valve 20 is slightly closed.

As described above, FIG. 8 shows the case where the fuel is burned under the stoichiometric air-fuel ratio. However, even if the air amount is made smaller than that shown in FIG. 8, that is, the air-fuel ratio is made rich. However, the amount of NO _x generated is 10 p.p. while preventing the generation of soot.
It can be around m or less, and even if the air amount is made larger than that shown in FIG. 8, that is, even if the average value of the air-fuel ratio is lean from 17 to 18, soot generation is prevented. Meanwhile, the amount of NO _x generated can be set to around 10 p.pm or less.

That is, when the air-fuel ratio is made rich, the fuel becomes excessive, but since the combustion temperature is suppressed to a low temperature, the excessive fuel does not grow to soot, and soot is generated. There is no. Further, at this time, a very small amount of NO _x is generated. On the other hand, when the average air-fuel ratio is lean, or even when the air-fuel ratio is the stoichiometric air-fuel ratio, a small amount of soot is generated if the combustion temperature becomes high, but in the present invention the combustion temperature is suppressed to a low temperature, soot Not generated at all. Furthermore, NO _x
Also produces only a very small amount.

As described above, when low temperature combustion is performed, soot is generated regardless of the air-fuel ratio, that is, whether the air-fuel ratio is rich, the stoichiometric air-fuel ratio, or the average air-fuel ratio is lean. However, the amount of NO _x generated is extremely small. Therefore, considering the improvement of the fuel consumption rate, it can be said that it is preferable to make the average air-fuel ratio lean at this time.

By the way, the temperature of the fuel and the gas around it in the combustion chamber during combustion can be suppressed to a temperature below the temperature at which the growth of hydrocarbons stops halfway only when the engine is operating at low load, where the calorific value of combustion is relatively small. To be Therefore, in the embodiment according to the present invention, when the engine is operated at low load, the temperature of the fuel and the gas around it during combustion is suppressed below the temperature at which the growth of hydrocarbons stops halfway, and the first combustion, that is, low temperature combustion is performed. In addition, the second combustion, that is, the combustion normally performed from the conventional one, is performed during the engine high load operation. It should be noted that here, the first combustion, that is, low temperature combustion, as is clear from the above description, the amount of inert gas in the combustion chamber is larger than the amount of inert gas at which the amount of soot generated peaks, and soot is almost generated. The second combustion, that is, the combustion that is normally performed in the past, is the combustion that does not have the amount of soot generated and the amount of the inert gas in the combustion chamber is smaller than the amount of the inert gas that does not reach the peak. Say that.

FIG. 9 shows a first operating region I in which the first combustion, that is, low temperature combustion is performed, and a second combustion region II in which the second combustion, that is, combustion by the conventional combustion method is performed. There is. In FIG. 9, the vertical axis TQ represents the required torque and the horizontal axis N represents the engine speed. Further, in FIG. 9, X (N) is the first operating region I and the second operating region II.
, And Y (N) indicates the second boundary between the first operating region I and the second operating region II.
The change determination of the operating region from the first operating region I to the second operating region II is performed based on the first boundary X (N),
The change determination of the operating region from the operating region II to the first operating region I is performed based on the second boundary Y (N).

That is, the operating condition of the engine is the first operating region I.
Required torque TQ when low temperature combustion is performed
Is above the first boundary X (N) which is a function of the engine speed N, it is judged that the operating region has moved to the second operating region II,
Combustion is performed by a conventional combustion method. Next, the required torque TQ is a second boundary Y (N) which is a function of the engine speed N.
When it becomes lower than that, it is determined that the operating region has moved to the first operating region I, and low temperature combustion is performed again.

In this way, the two boundaries of the first boundary X (N) and the second boundary Y (N) on the lower torque side of the first boundary X (N) are provided as follows. For one reason. The first reason is that the combustion temperature is relatively high on the high torque side of the second operating region II, and at this time the required torque TQ is the first boundary X.
This is because even if the temperature becomes lower than (N), low temperature combustion cannot be performed immediately. That is, the low temperature combustion is not started immediately unless the required torque TQ becomes considerably low, that is, when it becomes lower than the second boundary Y (N). The second reason is the first operating region I and the second operating region II.
This is because a hysteresis is provided for changes in the operating region between.

By the way, when the engine is operating in the first operating region I and low-temperature combustion is being carried out, soot is scarcely generated, and instead, unburned hydrocarbons are in a soot precursor or in a state before that. It is discharged from the combustion chamber 5 in shape. At this time, the unburned hydrocarbons discharged from the combustion chamber 5 are satisfactorily oxidized by the catalyst 25 having an oxidizing function. Catalyst 2
As 5, an oxidation catalyst, a three-way catalyst, or a NO _x absorbent can be used. _The NO _x absorbent absorbs NO _x when the mean air-fuel ratio in the combustion chamber 5 of the lean, the combustion chamber 5
It has a function of releasing NO _x when the average air-fuel ratio in the inside becomes rich.

This NO _x absorbent uses, for example, alumina as a carrier, and potassium K, sodium N, etc. are provided on this carrier.
a, at least one selected from alkali metals such as lithium Li and cesium Cs, alkaline earths such as barium Ba and calcium Ca, rare earths such as lanthanum La and yttrium Y, and a noble metal such as platinum Pt. Is carried.

Not only oxidation catalysts, but also three-way catalysts and NO
_{The x-} absorbent also has an oxidizing function, so that the three-way catalyst and the NO _x absorbent can be used as the catalyst 25 as described above. FIG. 10 shows the output of the air-fuel ratio sensor 27. As shown in FIG. 10, the output current I of the air-fuel ratio sensor 27 changes according to the air-fuel ratio A / F. Therefore, the air-fuel ratio can be known from the output current I of the air-fuel ratio sensor 27.

Next, referring to FIG. 11, the first operating region I
The operation control in the second operation area II will be briefly described. FIG. 11 shows the opening of the throttle valve 20, the opening of the EGR control valve 31, and the EGR with respect to the required torque TQ.
The ratio, the air-fuel ratio, the injection timing and the injection amount are shown. Figure 1
As shown in No. 1, in the first operating region I where the required torque TQ is low, the opening degree of the throttle valve 20 is gradually increased from near full closing to about 2/3 opening degree as the required load L increases, and the EGR The opening degree of the control valve 31 is gradually increased from near full close to full open as the required load L increases. Further, in the example shown in FIG. 11, the EGR rate is set to approximately 70% in the first operating region I, and the air-fuel ratio is set to a slightly lean lean air-fuel ratio.

In other words, in the first operating region I, EGR
The opening of the throttle valve 20 and the opening of the EGR control valve 31 are controlled so that the ratio becomes approximately 70% and the air-fuel ratio becomes a slightly lean air-fuel ratio. At this time, the air-fuel ratio is controlled to the target lean air-fuel ratio by correcting the opening degree of the EGR control valve 31 based on the output signal of the air-fuel ratio sensor 27. Further, in the first operation region I, fuel injection is performed before the compression top dead center TDC. In this case, the injection start timing θS becomes late as the required load L becomes high, and the injection completion timing θE also becomes late as the injection start timing θS becomes late.

During the idling operation, the throttle valve 20 is closed to the fully closed state, and the EGR control valve 31 is also closed to the fully closed state at this time. Throttle valve 2
When 0 is closed to near full closure, the pressure in the combustion chamber 5 at the beginning of compression becomes low and the compression pressure becomes small. When the compression pressure becomes small, the compression work by the piston 4 becomes small, so that the vibration of the engine body 1 becomes small. That is, during idling operation, the throttle valve 20 is closed close to the fully closed state in order to suppress the vibration of the engine body 1.

On the other hand, the engine operating condition is the first operating region I.
When changing from the second operating range II to the second operating range II, the opening degree of the throttle valve 20 is increased stepwise from about 2/3 opening degree toward the full opening direction. At this time, in the example shown in FIG. 11, the EGR rate is reduced stepwise from approximately 70% to 40% or less, and the air-fuel ratio is increased stepwise. That is, the EGR that produces a large amount of smoke with an EGR rate
The operating range of the engine is the first because the rate range (Fig. 7) is exceeded.
A large amount of smoke does not occur when changing from the operating region I to the second operating region II.

In the second operation region II, the second combustion, that is, the combustion which is conventionally performed is performed. Although soot and NO _x are slightly generated in this combustion method, the thermal efficiency is higher than that in low temperature combustion, and therefore, when the operating region of the engine changes from the first operating region I to the second operating region II, as shown in FIG. In addition, the injection amount is reduced stepwise. In this second operating region II, the throttle valve 20 is kept fully open except for a part, and the opening degree of the EGR control valve 31 is equal to the required torque TQ.
Becomes higher and smaller. Also, this operating area
In II, the EGR rate becomes lower as the required torque TQ becomes higher, and the air-fuel ratio becomes smaller as the required torque TQ becomes higher. However, the air-fuel ratio is set to the lean air-fuel ratio even if the required torque TQ becomes high. Further, in the second operation region II, the injection start timing θS is set near the compression top dead center TDC.

FIG. 12A shows the relationship between the required torque TQ, the depression amount L of the accelerator pedal 50, and the engine speed N. Note that each curve in FIG. 12A represents an equal torque curve, the curve indicated by TQ = 0 indicates that the torque is zero, and the remaining curves indicate TQ =.
The required torque gradually increases in the order of a, TQ = b, TQ = c, TQ = d. Required torque T shown in FIG.
Q is the accelerator pedal 50 as shown in FIG.
It is stored in advance in the ROM 42 in the form of a map as a function of the depression amount L and the engine speed N. In the embodiment according to the present invention, the accelerator pedal 5 is selected from the map shown in FIG.
The required torque TQ corresponding to the depression amount L of 0 and the engine speed N is first calculated, and the fuel injection amount and the like are calculated based on the required torque TQ.

FIG. 13 shows the air-fuel ratio A / F in the first operating region I. In FIG. 13, A / F = 15.
5, each curve indicated by A / F = 16, A / F = 17, A / F = 18 has an air-fuel ratio of 15.5, 16, 17, 18 respectively.
And the air-fuel ratio between the curves is determined by proportional distribution. As shown in FIG. 13, the air-fuel ratio is lean in the first operating region I, and in the first operating region I, the lower the required torque TQ is, the air-fuel ratio A /
F is lean.

That is, the lower the required torque TQ, the smaller the amount of heat generated by combustion. Therefore, as the required torque TQ becomes lower, the low temperature combustion can be performed even if the EGR rate is lowered. If the EGR rate is reduced, the air-fuel ratio will increase,
Therefore, as shown in FIG. 13, the air-fuel ratio A / F is increased as the required torque TQ decreases. Air-fuel ratio A / F
Becomes larger, the fuel consumption rate increases. Therefore, in order to make the air-fuel ratio as lean as possible, in the embodiment of the present invention, the air-fuel ratio A / F is made larger as the required torque TQ becomes lower.

[0069] FIG. 14 (A) shows the injection amount Q in the first operating region I, FIG. 14 (B) shows the reference injection start timing θS in the first operating region I. 14
As shown in (A), the injection amount Q in the first operating region I is stored in advance in the ROM 42 in the form of a map in the form of a map as a function of the required torque TQ and the engine speed N.
As shown in FIG. 4 (B), the reference injection start timing θS in the first operating region I is also stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N.

Further, the throttle valve 20 required to set the air-fuel ratio to the target air-fuel ratio A / F shown in FIG. 13 according to the operating state of the engine and to set the EGR rate to the target EGR rate according to the operating state of the engine. The target opening degree ST is stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N as shown in FIG. 15 (A), and the air-fuel ratio is shown according to the operating state of the engine. Target air-fuel ratio A / shown in 13
The target EG is set to F and the EGR rate is set according to the operating state of the engine.
The target opening degree SE of the EGR control valve 31 required to obtain the R rate
Is a ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N as shown in FIG.
It is stored in.

Further, in the present invention, the air-fuel ratio is set to the target air-fuel ratio A / F shown in FIG.
The pressure PM in the air suction pipe 17 downstream of the throttle valve 20 when the rate is set to the target EGR rate according to the operating state of the engine
0 is a function of the required torque TQ and the engine speed N as shown in FIG.
It is stored in 2.

FIG. 16 shows the target air-fuel ratio when the second combustion, that is, the normal combustion by the conventional combustion method is performed. In FIG. 16, A / F = 24, A / F = 3
5, each curve shown by A / F = 45 and A / F = 60 shows the target air-fuel ratios 24, 35, 45, 60, respectively. 17A shows the injection amount Q in the second operation region II, and FIG. 17B shows the injection start timing θS in the second operation region II. As shown in FIG. 17 (A), the injection amount Q in the second operation region II is the required torque T
R in advance in the form of a map as a function of Q and engine speed N
It is stored in the OM 42, and as shown in FIG. 17B, the injection start timing θS in the second operation region II is also stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N. Has been done.

Further, the target opening degree ST of the throttle valve 20 required to bring the air-fuel ratio to the target air-fuel ratio shown in FIG.
As shown in FIG. 8 (A), the EGR is stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N, and is required to set the air-fuel ratio to the target air-fuel ratio shown in FIG. The target opening degree SE of the control valve 31 is shown in FIG.
As shown in (B), it is stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N.

Now, in the first operating region I, the injection amount is the injection amount Q calculated from the map shown in FIG. 14 (A), and the opening of the throttle valve 20 is the target opening shown in FIG. 15 (A). ST and the opening degree of the EGR control valve 31 is shown in FIG.
When the target opening degree SE shown in Fig.
The target air-fuel ratio A / F becomes as shown in, and the target EGR according to the required torque TQ and the engine speed N at that time is the EGR rate.
Become a rate. In addition, the injection amount is the injection amount Q shown in FIG.
The target air-fuel ratio A / F shown in FIG.
Therefore, the intake air amount at this time is the target intake air amount corresponding to the required torque TQ and the engine speed N at this time. At this time, the EGR gas amount is also the target EGR gas amount corresponding to the required torque TQ and the engine speed N at this time.

By the way, the pressure in the air suction pipe 17 downstream of the throttle valve 20 is equal to that of the air suction pipe 1 downstream of the throttle valve 20.
It is determined by the amount of intake air and the amount of EGR gas flowing into 7. Therefore, as described above, when the intake air amount and the EGR gas amount are the target values, respectively, the pressure in the air suction pipe 17 downstream of the throttle valve 20 is a pressure corresponding to these target values, and the pressure at this time is It matches the target pressure PM0 shown in FIG. 15 (C) corresponding to the required torque TQ and the engine speed N.

However, in reality, the dimensional variation of parts, the secular change, the throttle valve 20 or the EGR control valve 3
Even if the injection amount, the opening degree of the throttle valve 20 and the opening degree of the EGR control valve 31 are determined based on the corresponding map due to the clogging of 1, the air-fuel ratio does not exactly match the air-fuel ratio shown in FIG. The rate deviates from the target EGR rate. Therefore, in the first embodiment according to the present invention, the target intake air amount required to bring the air-fuel ratio to the target air-fuel ratio A / F is calculated from the target injection amount Q, and the mass of the intake air detected by the mass flow rate detector 21 is calculated. The opening degree of the throttle valve 20 is corrected so that the flow rate (hereinafter, simply referred to as the intake air amount) becomes the target intake air amount, so that the air-fuel ratio accurately matches the target air-fuel ratio.

As described above, in the first embodiment, the air-fuel ratio is made to exactly match the target air-fuel ratio, that is, the intake air amount is made to exactly match the target intake air amount. If it matches the EGR gas amount, the pressure in the air suction pipe 17 downstream of the throttle valve 20 will match the target pressure PM0 shown in FIG. 15 (C). In other words, at this time, when the pressure in the air suction pipe 17 downstream of the throttle valve 20 deviates from the target pressure PM0 shown in FIG. 15C, the EGR gas amount is the target E.
It does not match the GR gas amount, and thus the EGR rate does not match the target EGR rate.

Therefore, in the embodiment according to the present invention, the pressure downstream of the throttle valve 20 is the target pressure P shown in FIG.
When it deviates from M0, the pressure downstream of the throttle valve 20 is adjusted to the target pressure shown in FIG.
The opening degree of the R control valve 31 is controlled so that the EGR rate matches the target EGR rate. Also, the first
In the embodiment, the engine speed is controlled so that the engine speed becomes the target idling speed during engine idling operation. In the first embodiment, the control of the engine speed is performed by controlling the fuel injection amount. Even in this case, the air-fuel ratio is controlled so as to reach the target air-fuel ratio. Also,
The opening degree of the EGR control valve 31 is controlled so that the pressure in the air suction pipe 17 downstream of the throttle valve 20 becomes the target pressure,
Thus, the EGR rate is controlled so as to become the target EGR rate.

In this case, maintaining the pressure in the air suction pipe 17 downstream of the throttle valve 20 at the target pressure has two meanings. One is to ensure good low temperature combustion by controlling the EGR rate to the target EGR rate. The other is to suppress the vibration of the engine body 1 by suppressing the pressure in the combustion chamber 5 at the beginning of compression to be low. Next, the operation control will be described with reference to FIG.

Referring to FIG. 19, first, at step 100, it is judged if the flag I indicating that the engine operating condition is the first operating region I is set or not. When the flag I is set, that is, when the operating region of the engine is the first operating region I, step 1
The routine proceeds to 01 where it is determined whether the required torque TQ has become larger than the first boundary X1 (N). TQ ≤ X1
In the case of (N), the routine proceeds to step 103, where the operation control I for executing the first combustion is performed. A routine for executing this operation control I is shown in FIGS. 20 and 21.

On the other hand, in step 101, TQ> X
When it is determined that (N) has been reached, the routine proceeds to step 102, where the flag I is reset, and then step 106.
Then, the operation control II for executing the second combustion is performed. The routine for executing this operation control II is shown in FIG.
7 is shown. When the flag I is reset, in the next processing cycle, the routine proceeds from step 100 to step 104, where it is judged if the required torque TQ has become lower than the second boundary Y (N). When TQ ≧ Y (N), the routine proceeds to step 106, where the second combustion is performed. On the other hand, when it is determined in step 104 that TQ <Y (N), the routine proceeds to step 105, where the flag I is set, then the routine proceeds to step 103, where low temperature combustion is performed.

Next, the operation control I for executing the low temperature combustion will be described with reference to FIGS . 20 and 21 . Referring to FIGS. 20 and 21 , first, step 20
0, the target opening degree ST of the throttle valve 20 is calculated from the map shown in FIG.
In FIG. 15B, the map shown in FIG.
The target opening degree SE of 1 is calculated. Then step 202
Then, the injection amount Q is calculated from the map shown in FIG. Next, at step 203, it is judged if the engine is idling or not. For example, accelerator pedal 50
It is determined that the engine is idling when the depression amount of the vehicle is zero and the vehicle speed is zero.

When the engine is not idling, the routine jumps to step 208 and the target air-fuel ratio t shown in FIG.
(A / F) is calculated. Next, at step 209, the target intake air amount tGa required to bring the air-fuel ratio to the target air-fuel ratio t (A / F) is calculated based on the injection amount Q and the target air-fuel ratio t (A / F). Next, at step 210, the actual intake air amount G detected by the mass flow rate detector 21.
a is captured. Next, at step 211, it is judged if the actual intake air amount Ga is larger than the target intake air amount tGa.

When Ga> tGa, the routine proceeds to step 212, where the constant value a is subtracted from the correction amount ΔST for the throttle valve opening, and then the routine proceeds to step 214. On the other hand, when Ga ≦ tGa, the routine proceeds to step 213, where a constant value a is added to the correction amount ΔST, and then step 2
Proceed to 14. In step 214, a value obtained by adding the correction value ΔST to the target opening degree ST of the throttle valve 20 (= ST + ΔS
T) is the final opening ST of the throttle valve 20.
Therefore, when Ga> tGa, the opening degree of the throttle valve 20 is decreased, and when Ga ≦ tGa, the opening degree of the throttle valve 20 is increased, whereby the actual intake air amount Ga becomes the target intake air amount tGa, The air-fuel ratio is set to the target air-fuel ratio t (A / F).

In step 215, the target pressure PM0 in the air suction pipe 17 downstream of the throttle valve 20 is calculated from the map shown in FIG. 15 (C). Then the pressure PM in the air suction pipe 17 detected by the pressure sensor 3 8 at step 216 whether or not higher than the target pressure PM0 is determined. When PM> PM0, the process proceeds to step 217 and E
The constant value b is subtracted from the correction value ΔSE for the GR control valve 31, and then the routine proceeds to step 219. On the other hand, P
When M ≦ PM0, the routine proceeds to step 218, where the correction value Δ
The constant value b is added to SE, and then the process proceeds to step 219.

At step 219, a value (= SE + ΔS) obtained by adding the correction value ΔSE to the target opening degree SE of the EGR control valve 31.
E) is the final opening SE of the EGR control valve 31.
Therefore, when PM> PM0, the opening degree of the EGR control valve 31 is reduced, and when PM ≦ PM0, the opening degree of the EGR control valve 31 is increased, whereby the EGR rate becomes the target EGR rate. Next, the routine proceeds to the injection timing control routine shown in FIG.

On the other hand, if it is determined in step 203 that the engine is idling, then step 20
4, the engine speed N is set to the target idling speed N _0.
Is higher than the above. When N> N _0, the routine proceeds to step 205, where the constant value C is subtracted from the correction value ΔQ for the injection amount, and then the routine proceeds to step 207. On the other hand, when N ≦ N _0, the routine proceeds to step 206, where the constant value C is added to the correction value ΔQ, and then step 207
Proceed to.

In step 207, the final injection amount Q is the value (= Q + ΔQ) obtained by adding the correction value ΔQ to the injection amount Q calculated from the map. Therefore, when N> N ₀ , the injection amount is decreased, and when N ≦ N ₀ , the injection amount is increased, whereby the engine speed N is set to the target idling speed N ₀ . Next, in steps 208 to 214, the intake air amount is made the target intake air amount, and the air-fuel ratio is made the target air-fuel ratio t (A / F).
Next, in steps 216 to 219, the pressure PM downstream of the throttle valve 20 is set to the target pressure PM0. At this time, the EGR rate becomes the target EGR rate.

Before describing the injection timing control routine shown in FIG. 25, the injection timing control method will be described with reference to FIG. Whether in an embodiment according to the present invention has been made good low temperature combustion is determined based on the pressure in the combustion chamber 5 detected by the combustion pressure sensor 3 7. That is,
When good low temperature combustion is being performed, the combustion pressure changes gently as shown in FIG. Specifically, the combustion pressure once peaks at the top dead center TDC as indicated by P ₀ , and then peaks again after the top dead center TDC as indicated by P ₁ . The peak pressure P ₁ is generated by the combustion pressure, and when good low temperature combustion is being performed, the increase amount of the peak pressure P ₁ with respect to the peak pressure P ₀ , that is, the differential pressure ΔP (= P difference between the peak pressure P ₀ and the peak pressure P ₁ = P ₁ -P
₀ ) becomes relatively small.

On the other hand, for example, a region where the density of fuel particles is high is locally formed, and as a result, when the amount of pressure increase after ignition increases, the combustion temperature increases. At this time, low temperature combustion is no longer performed, and thus a large amount of soot is generated. Therefore, in the embodiment according to the present invention, the differential pressure ΔP (= P
_{When 1−} P ₀ ) exceeds a predetermined upper limit value α, the injection timing is delayed so that the differential pressure ΔP becomes smaller.

As shown in FIG. 23A, the upper limit value α becomes smaller as the required torque TQ becomes larger.
As shown in (B), the upper limit α decreases as the engine speed N increases. This upper limit value α is stored in advance in the ROM 42 in the form of a map as a function of the required torque TQ and the engine speed N, as shown in FIG.

Further, if low temperature combustion is not performed satisfactorily and combustion failure occurs, the peak pressure P ₁ becomes lower than the peak pressure P ₀ . Therefore, in the embodiment according to the present invention, the differential pressure ΔP (= P ₁ −
When P ₀ ) becomes negative, the injection timing is advanced so that good low temperature combustion is performed. Next, a method of detecting the differential pressure ΔP will be described with reference to FIGS. 22 and 24. FIG. 24 shows a crank angle interrupt routine,
First, at step 300, the current crank angle is C
It is determined whether or not it is A1 (FIG. 22). When the crank angle is CA1, the routine proceeds to step 301, where the output voltage of the peak hold circuit 49 is read. At this time, the output voltage of the peak hold circuit 49 represents the peak pressure P _0. Therefore, in step 301, the peak pressure P ₀ is read. Next, at step 302, the reset signal is input to the reset input terminal R of the peak hold circuit 49, whereby the peak hold circuit 49 is reset.

Next, at step 303, it is judged if the present crank angle is CA2 (FIG. 22). When the crank angle is CA2, the routine proceeds to step 304, where the output voltage of the peak hold circuit 49 is read. At this time, the output voltage of the peak hold circuit 49 represents the peak pressure P ₁ , and therefore the peak pressure P ₁ is read in step 304. Next, at step 305, the reset signal is input to the reset input terminal R of the peak hold circuit 49, whereby the peak hold circuit 49 is reset. Next, at step 306, the differential pressure ΔP (= P ₁ −P ₀ ) between the peak pressure P ₀ and the peak pressure P ₁ is calculated.

Next, the injection timing control routine shown in FIG. 25 will be described. Referring to FIG. 25, first, at step 400, the reference injection start timing θS is calculated from the map shown in FIG. 14 (B). Then step 40
At 1, it is determined whether the differential pressure ΔP (= P ₁ −P ₀ ) is greater than zero. When ΔP ≧ 0, the routine proceeds to step 406, where the upper limit value α is calculated from the map shown in FIG. Next, at step 407, it is judged if the differential pressure ΔP is smaller than the upper limit value α. When ΔP <α, the processing cycle is completed. That is, when 0 ≦ ΔP <α, the processing cycle is completed.

On the other hand, when it is judged at step 407 that ΔP ≧ α, the routine proceeds to step 408, where a constant value e is added to the correction value Δθ for the reference injection start timing θS. Next, at step 409, the correction value Δθ is subtracted from the reference injection start timing θS, whereby the injection start timing θ
S is delayed. Next, at step 410, the maximum allowable retard angle θ _min is calculated. This maximum allowable retard angle θ
_As shown in FIG. 26, _min is stored in advance in the ROM 42 as a function of the required torque TQ and the engine speed N. Then whether step 411, the injection start timing θS becomes later than the allowable maximum retard timing theta _{mi n,} i.e. theta
It is determined whether or not S <θ _min . If θS ≧ θ _min , the processing cycle is completed. On the other hand, θS <θ
_{When it reaches min} , the routine proceeds to step 412, where the injection start timing θS
Is the maximum allowable retard angle θ _min .

On the other hand, when it is judged in step 401 that the differential pressure ΔP is negative, the routine proceeds to step 402, where a constant value e is subtracted from the correction value Δθ. Then step 403
Then, the correction value Δθ is subtracted from the reference injection start timing θS,
At this time, the injection start timing θS is advanced. Next, at step 404, it is judged if the correction value Δθ is larger than zero. When Δθ ≧ 0, the processing cycle is completed.
On the other hand, when ΔQ <0, the routine proceeds to step 405, where the injection start timing θS is made the reference injection start timing obtained from the map of FIG. 14 (B).

Thus, the throttle valve 2 should be used for low temperature combustion.
When the opening degree of 0 and the opening degree of the EGR control valve 31 are controlled, if the differential pressure ΔP becomes larger than the upper limit value α, the injection start timing is gradually delayed, and if the differential pressure ΔP becomes negative, the injection start timing is increased. Is gradually accelerated. As a result, good low temperature combustion can always be performed. Next, referring to FIG. 27, FIG.
The routine of the operation control II for executing the second combustion performed in step 106 of is described.

Referring to FIG. 27, first, step 5
At 00, the target fuel injection amount Q is calculated from the map shown in FIG. 17 (A), and the fuel injection amount is the target fuel injection amount Q.
It is said that Next, at step 501, the target opening degree ST of the throttle valve 20 is calculated from the map shown in FIG. Next, at step 502, the target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG.
The opening degree of the EGR control valve 31 is set to this target opening degree SE.

Next, at step 503, the intake air amount Ga detected by the mass flow detector 21 is taken in. Next, at step 504, the fuel injection amount Q and the intake air amount Ga
From this, the actual air-fuel ratio A / F is calculated. Next, at step 605, the target air-fuel ratio t (A / F) shown in FIG. 16 is calculated. Next, at step 606, the actual air-fuel ratio A / F
Is larger than the target air-fuel ratio t (A / F). When A / F> t (A / F), step 507
Then, the correction value ΔST of the throttle opening is reduced by a constant value α, and then the routine proceeds to step 509. Step 5 08 when contrast of A / F ≦ t (A / F)
And the correction value ΔST is increased by a constant value α.
Then, it proceeds to step 509. In step 509, the final target opening ST is calculated by adding the correction value ΔST to the target opening ST of the throttle valve 20. That is,
The opening of the throttle valve 20 is controlled so that the actual air-fuel ratio A / F becomes the target air-fuel ratio t (A / F). Next, at step 510, the injection start timing θS is calculated from the map shown in FIG. 17 (B).

FIG. 28 shows another embodiment for executing the operation control II. Referring to FIG. 28, first, at step 600, the target fuel injection amount Q is calculated from the map shown in FIG. 17 (A), and the fuel injection amount is made this target fuel injection amount Q. Next, in step 601, FIG.
From the map shown in (A), the target opening S of the throttle valve 20
T is calculated, and the opening of the throttle valve 20 is set to this target opening ST. Next, in step 602, FIG.
The target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG.

Next, at step 603, the intake air amount Ga detected by the mass flow rate detector 21 is taken in. Next, at step 604, the fuel injection amount Q and the intake air amount Ga
From this, the actual air-fuel ratio A / F is calculated. Next, at step 605, the target air-fuel ratio t (A / F) shown in FIG. 16 is calculated. Next, at step 606, the actual air-fuel ratio A / F
Is larger than the target air-fuel ratio t (A / F). When A / F> t (A / F), step 607
Then, the correction value ΔSE for the opening degree of the EGR control valve is increased by a constant value α, and then the routine proceeds to step 609. On the other hand, when A / F ≦ t (A / F), the routine proceeds to step 608, where the correction value ΔSE is decreased by a constant value α, and then the routine proceeds to step 609. Step 60
At 9, the correction value ΔSE is added to the target opening degree SE of the EGR control valve 31.
The final target opening degree SE is calculated by adding That is, the actual air-fuel ratio A / F is the target air-fuel ratio t (A /
The opening degree of the EGR control valve 31 is controlled so as to be F). Next, at step 610, the injection start timing θS is calculated from the map shown in FIG. 17 (B).

29 and 30 show the second embodiment of the operation control I. In this embodiment, the intake air amount is controlled so that the actual air-fuel ratio detected by the air-fuel ratio sensor 27 becomes the target air-fuel ratio during low temperature combustion. Referring to FIG. 29 and FIG. 30, first, at step 700, the target opening degree ST of the throttle valve 20 is calculated from the map shown in FIG.
The target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG. Next, in step 702, FIG.
The injection amount Q is calculated from the map shown in (A). Next, at step 703, it is judged if the engine is idling or not. As described above, for example, when the amount of depression of the accelerator pedal 50 is zero and the vehicle speed is zero, it is determined that the engine is idling.

When the engine is not idling, the routine jumps to step 708 and the target air-fuel ratio t shown in FIG.
(A / F) is calculated. Next, at step 709, it is judged if the actual air-fuel ratio A / F detected by the air-fuel ratio sensor 27 is larger than the target air-fuel ratio t (A / F). When A / F> t (A / F), the routine proceeds to step 710, where the constant value a is subtracted from the correction amount ΔST with respect to the throttle valve opening, and then the routine proceeds to step 712. On the other hand, when A / F ≦ t (A / F), step 711
Then, a constant value a is added to the correction amount ΔST, and then the process proceeds to step 712. In step 712, a value obtained by adding the correction value ΔST to the target opening degree ST of the throttle valve 20 (= S
T + ΔST) is the final opening ST of the throttle valve 20. Therefore, when A / F> t (A / F), the opening degree of the throttle valve 20 is reduced, and A / F ≦ t (A / F
In the case of F), the opening degree of the throttle valve 20 is increased so that the actual intake air amount Ga becomes the target intake air amount tGa and the air-fuel ratio becomes the target air-fuel ratio t (A / F).

In step 714, the target pressure PM0 in the air suction pipe 17 downstream of the throttle valve 20 is calculated from the map shown in FIG. 15 (C). Then the pressure PM in the air suction pipe 17 detected by the pressure sensor 3 8 at step 715 whether or not higher than the target pressure PM0 is determined. When PM> PM0, the routine proceeds to step 715, where E
The constant value b is subtracted from the correction value ΔSE for the GR control valve 31, and then the process proceeds to step 717. On the other hand, P
When M ≦ PM0, the routine proceeds to step 716, where the correction value Δ
The constant value b is added to SE, and then the process proceeds to step 717.

At step 717, the value obtained by adding the correction value ΔSE to the target opening degree SE of the EGR control valve 31 (= SE + ΔS
E) is the final opening SE of the EGR control valve 31.
Therefore, when PM> PM0, the opening degree of the EGR control valve 31 is reduced, and when PM ≦ PM0, the opening degree of the EGR control valve 31 is increased, whereby the EGR rate becomes the target EGR rate. Next, the routine proceeds to the injection timing control routine shown in FIG.

On the other hand, if it is determined in step 703 that the engine is idling, then step 70
4, the engine speed N is set to the target idling speed N _0.
Is higher than the above. When N> N _0, the routine proceeds to step 705, where the constant value C is subtracted from the correction value ΔQ for the injection amount, and then the routine proceeds to step 707. On the other hand, when N ≦ N _0, the routine proceeds to step 706, where the constant value C is added to the correction value ΔQ, and then step 707.
Proceed to.

At step 707, the final injection amount Q is the value (= Q + ΔQ) obtained by adding the correction value ΔQ to the injection amount Q calculated from the map. Therefore, when N> N ₀ , the injection amount is decreased, and when N ≦ N ₀ , the injection amount is increased, whereby the engine speed N is set to the target idling speed N ₀ . Next, at steps 708 to 712, the intake air amount is made the target intake air amount and the air-fuel ratio is made the target air-fuel ratio t (A / F).
Next, in steps 714 to 717, the pressure PM downstream of the throttle valve 20 is set to the target pressure PM0. At this time, the EGR rate becomes the target EGR rate.

A third embodiment of the operation control I is shown in FIGS. 31 to 35. In this embodiment, the target air-fuel ratio A / F is predetermined in the form of a function of the intake air amount Ga and the engine speed N as shown in FIG. 31, and the actual intake air detected by the mass flow rate detector 21 is detected. The fuel injection amount Q required to bring the air-fuel ratio to the target air-fuel ratio A / F is calculated based on the air amount Ga and the target air-fuel ratio A / F.

Further, in this embodiment, the target pressure PM0 in the air suction pipe 17 downstream of the throttle valve 20 is also shown in FIG.
As shown in Fig. 33, the intake air amount Ga and the engine speed N are pre-stored in the ROM 42 in the form of a map as a function, and the injection start timing &thgr; S is also shown in Fig. 33 (B). It is previously stored in the ROM 42 in the form of a map as a function of the engine speed N. The target pressure PM0 and the injection start timing θS are the injection amount Q
And RO in advance in the form of a map as a function of the engine speed N
It may be stored in M42.

Further, in this embodiment, the engine speed is controlled to the target idling speed by controlling the opening degree of the throttle valve 20 during engine idling operation, that is, by controlling the intake air amount. Next, the operation control I will be described with reference to FIGS. 34 and 35.
Referring to FIGS. 34 and 35, first, step 8
At 00, the target opening degree ST of the throttle valve 20 is calculated from the map shown in FIG.
1, the target opening degree SE of the EGR control valve 31 is calculated from the map shown in FIG. Then step 80
In 2, it is determined whether or not the engine is idling. As described above, for example, when the amount of depression of the accelerator pedal 50 is zero and the vehicle speed is zero, it is determined that the engine is idling.

When the engine is not idling, the routine jumps to step 807 and the actual intake air amount Ga detected by the mass flow rate detector 21 is taken in. Next, at step 808, the target air-fuel ratio t (A
/ F) is calculated. Next, at step 809, the fuel injection amount Q required to bring the air-fuel ratio to the target air-fuel ratio t (A / F) based on the intake air amount Ga and the target air-fuel ratio t (A / F).
Is calculated.

Next, at step 810, from the map shown in FIG. 33A, the air suction pipe 17 downstream of the throttle valve 20 is drawn.
The target pressure PM0 therein is calculated. Then step 81
1 air suction pipe 17 detected by the pressure sensor 3 8,
It is determined whether the internal pressure PM is higher than the target pressure PM0. When PM> PM0, the routine proceeds to step 812, where the correction value ΔSE for the EGR control valve 31 is changed to a constant value b.
Is subtracted, then step 814 is reached. On the other hand, when PM ≦ PM0, the routine proceeds to step 813, where the constant value b is added to the correction value ΔSE, and then step 81
Go to 4.

At step 814, a value obtained by adding the correction value ΔSE to the target opening degree SE of the EGR control valve 31 (= SE + ΔS
E) is the final opening SE of the EGR control valve 31.
Therefore, when PM> PM0, the opening degree of the EGR control valve 31 is reduced, and when PM ≦ PM0, the opening degree of the EGR control valve 31 is increased, whereby the EGR rate becomes the target EGR rate. Next, the routine proceeds to the injection timing control routine shown in FIG.

On the other hand, if it is determined in step 802 that the engine is idling, then step 80
3, the engine speed N is set to the target idling speed N _0.
Is higher than the above. When N> N _0, the routine proceeds to step 804, where the constant value a is subtracted from the correction value ΔST for the throttle valve opening, and then step 806.
Proceed to. On the other hand, when N ≦ N ₀ , step 80
5, the constant value a is added to the correction value ΔST, and then the process proceeds to step 806.

In step 806, a value obtained by adding the correction value ΔST to the target opening degree ST of the throttle valve 20 (= ST + ΔS
T) is the final opening ST of the throttle valve 20.
Therefore, when N> N ₀ , the opening degree of the throttle valve 20 is decreased, and when N ≦ N ₀ , the opening degree of the throttle valve 20 is increased, whereby the engine speed N is made the target idling speed N _0. It Then step 8
From 07 to step 809, the air-fuel ratio is the target air-fuel ratio t
(A / F), and then in steps 810 to 814, the pressure PM downstream of the throttle valve 20 is set to the target pressure PM0. At this time, the EGR rate becomes the target EGR rate.

In the first embodiment (FIGS. 20 and 21) and the second embodiment (FIGS. 29 and 30) of the operation control I, the injection amount Q and the target air-fuel ratio t are calculated based on the required torque TQ and the engine speed N. (A / F) is calculated, and the intake air amount Ga is controlled based on the injection amount Q and the target air-fuel ratio t (A / F) so that the air-fuel ratio becomes the target air-fuel ratio t (A / F). . Therefore, in this case, as long as the actual injection amount matches the calculated injection amount Q, the intake air amount Ga becomes the target intake air amount according to the required torque TQ and the engine speed N.

However, in the third embodiment of the operation control I (FIGS. 34 and 35), the intake air amount is controlled only on the basis of the throttle valve opening and the EGR control valve opening. When clogging occurs, the actual intake air amount Ga deviates from the target intake air amount. In the third embodiment, even if the actual intake air amount Ga deviates from the target intake air amount as described above, the target air-fuel ratio becomes the target air-fuel ratio t (A / F) and the EGR rate becomes the target EGR rate. The fuel ratio is determined based on the intake air amount Ga and the engine speed N, and the target pressure PM downstream of the throttle valve 20 is set.
0 is determined based on the intake air amount Ga and the engine speed N.

The throttle valve 20 may be arranged downstream of the compressor 16 of the exhaust turbocharger 15 and the EGR passage 29 may be connected to the intake passage downstream of the throttle valve 20. The pressure sensor 3 8 is disposed in the throttle valve 20 downstream of the intake passage in this case, the throttle valve 2
The opening degree of the EGR control valve 31 is controlled so that the pressure in the intake passage downstream of 0 becomes the target pressure PM0.

[0119]

The EGR rate can be exactly matched with the target EGR rate.

[Brief description of drawings]

FIG. 1 is an overall view of a compression ignition type internal combustion engine.

FIG. 2 is an enlarged side sectional view around a throttle valve.

FIG. 3 is an enlarged side sectional view around a throttle valve showing another embodiment.

FIG. 4 is a diagram showing amounts of smoke and NO _x generated, and the like.

FIG. 5 is a diagram showing a combustion pressure.

FIG. 6 is a diagram showing a fuel molecule.

FIG. 7 is a diagram showing the relationship between the amount of smoke generated and the EGR rate.

FIG. 8 is a diagram showing a relationship between an injected fuel amount and a mixed gas amount.

FIG. 9 is a diagram showing a first operating region I and a second operating region II.

FIG. 10 is a diagram showing an output of an air-fuel ratio sensor.

FIG. 11 is a diagram showing an opening of a throttle valve and the like.

FIG. 12 is a diagram showing a required torque.

13 is a diagram showing an air-fuel ratio in a first operating region I. FIG.

FIG. 14 is a diagram showing a map of an injection amount and the like.

FIG. 15 is a diagram showing a map of a target opening degree etc. of a throttle valve.

FIG. 16 is a diagram showing an air-fuel ratio in the second combustion.

FIG. 17 is a diagram showing a map of an injection amount and the like.

FIG. 18 is a diagram showing a map of a target opening degree etc. of a throttle valve.

FIG. 19 is a flowchart for controlling the operation of the engine.

FIG. 20 is a flowchart showing a first embodiment for executing operation control I.

FIG. 21 is a flowchart showing a first embodiment for executing operation control I.

FIG. 22 is a diagram showing combustion pressure and the like.

FIG. 23 is a diagram showing an upper limit value α.

FIG. 24 is a diagram showing a crank angle interruption routine.

FIG. 25 is a flowchart for controlling the injection timing.

FIG. 26 is a diagram showing a map of an allowable maximum retard timing.

FIG. 27 is a flowchart for executing operation control II.

FIG. 28 is a flowchart showing another embodiment for executing operation control II.

FIG. 29 is a flowchart showing a second embodiment for executing the operation control I.

FIG. 30 is a flowchart showing a second embodiment for executing operation control I.

FIG. 31 is a diagram showing an air-fuel ratio in a first operating region I.

FIG. 32 is a diagram showing a map of a target opening degree of a throttle valve or the like.

FIG. 33 is a diagram showing a map of target pressure and the like in the intake passage downstream of the throttle valve.

FIG. 34 is a flowchart showing a third embodiment for executing operation control I.

FIG. 35 is a flowchart showing a third embodiment for executing operation control I.

[Explanation of symbols]

6 ... Fuel injection valve 20 ... Throttle valve 31 ... EGR control valve 37 ... Pressure sensor

─────────────────────────────────────────────────── ─── Continuation of front page (51) Int.Cl. ⁷ Identification code FI F01N 3/24 F01N 3/24 S F02D 21/08 301 F02D 21/08 301B 301D 41/02 351 41/02 351 41/04 360 41/04 360C 385 385J 41/14 310 41/14 310N 41/16 41/16 Z 45/00 301 45/00 301F (72) Inventor Takekazu Ito 1 Toyota-cho, Toyota-shi, Aichi Toyota Motor Corporation (72) Inventor Hiroki Murata 1 Toyota-cho, Toyota-shi, Aichi Toyota Motor Co., Ltd. (72) Inventor Tsukasa Abe 1 Toyota-cho, Toyota-shi, Aichi Toyota Motor Co., Ltd. (56) References JP 7-4287 (JP, A) JP 8-254134 (JP, A) JP 9-32650 (JP, A) JP 9-79091 (JP, A) JP 8 -61112 (JP, A) JP-A-9-14016 (JP, A) JP-A-5-1628 (JP, A) JP-A-7-269375 (JP, A) JP-A-8-246935 (JP, A) ) JP-A-8-177654 (JP, A) JP-A-8-86251 (JP, A) JP-A-9-287527 (JP, A) (58) Fields investigated (Int.Cl. ⁷ , DB name) F02M 25/07 550 F02M 25/07 570 F01N 3/08 F01N 3/24 F02D 21/08 301 F02D 41/02 351 F02D 41/04 360 F02D 41/04 385 F02D 41/14 310 F02D 41/16 F02D 45 / 00 301

Claims (19)

(57) [Claims] 1. A recirculated exhaust gas generation amount of As you increase the recirculated exhaust gas amount supplied to the combustion chamber soot is <br/> peaked gradually increased, is supplied to the combustion chamber Further increase the amount
The fuel used during combustion in the combustion chamber and its
The ambient gas temperature becomes lower than the soot formation temperature,
In an internal combustion engine that rarely occurs, the amount of recirculated exhaust gas supplied to the combustion chamber is larger than the amount of recirculated exhaust gas that supplies the recirculated exhaust gas into the engine intake passage and the peak amount of soot is generated. In the combustion chamber by
The soot generates the temperature of the fuel and the gas around it during combustion.
The intake air amount control means for controlling the intake air amount supplied to the combustion chamber to a predetermined intake air amount according to the operating state of the engine by suppressing the temperature to a temperature lower than A pressure detecting means for detecting the pressure, and a recirculation exhaust gas amount control means for controlling the recirculation exhaust gas amount so that the exhaust gas recirculation rate becomes the target exhaust gas recirculation rate based on the pressure in the intake passage. Internal combustion engine equipped. 2. The recirculation exhaust gas is provided with a storage means for pre-storing a target pressure in the intake passage according to an engine operating condition necessary for making the exhaust gas recirculation rate the target exhaust gas recirculation rate. The internal combustion engine according to claim 1, wherein the amount control means controls the recirculated exhaust gas amount so that the pressure in the intake passage becomes the target pressure. 3. The recirculation exhaust gas amount control means comprises an exhaust gas recirculation control valve arranged in an exhaust gas recirculation passage connecting the engine exhaust passage and the engine intake passage, and the recirculation exhaust gas. The internal combustion engine according to claim 2, wherein the amount control means controls the opening degree of the exhaust gas recirculation control valve so that the pressure in the intake passage becomes the target pressure. 4. An intake air amount detection means for detecting the intake air amount, and an air-fuel ratio control means for controlling the air-fuel ratio to a target air-fuel ratio according to the operating state of the engine based on the detected intake air amount. The internal combustion engine according to claim 1, further comprising: 5. An air-fuel ratio control means is provided with a calculating means for calculating a fuel injection quantity, and the air-fuel ratio control means makes the ratio of the detected intake air quantity and the calculated fuel injection quantity a target air-fuel ratio. The internal combustion engine according to claim 4, which corrects 6. The air-fuel ratio detecting means for detecting an air-fuel ratio is provided, and the air-fuel ratio control means corrects the intake air amount so that the detected air-fuel ratio becomes a target air-fuel ratio. Internal combustion engine. 7. The internal combustion engine according to claim 4, wherein the air-fuel ratio control means calculates a fuel injection amount required to bring the air-fuel ratio to a target air-fuel ratio based on the detected intake air amount. 8. The intake air amount control means comprises a throttle valve disposed in the engine intake passage, and the intake air amount supplied into the combustion chamber is a predetermined intake air amount according to the operating state of the engine. The internal combustion engine according to claim 1, wherein the opening degree of the throttle valve is controlled so that. 9. A storage means for storing the injection timing is provided,
The internal combustion engine according to claim 1, wherein the injection timing is controlled based on the stored injection timing. 10. The internal combustion engine according to claim 9, further comprising combustion pressure detecting means for detecting the combustion pressure in the combustion chamber, and controlling the injection timing based on the combustion pressure. 11. The internal combustion engine according to claim 1, further comprising rotation speed control means for controlling the engine rotation speed to a target idling rotation speed during engine idling operation. 12. The internal combustion engine according to claim 11, wherein the engine speed control means maintains the engine speed at the target idling speed by controlling the fuel injection amount. 13. The internal combustion engine according to claim 11, wherein the engine speed control means maintains the engine speed at the target idling speed by controlling the intake air amount. 14. The intake air amount control means comprises a throttle valve arranged in the engine intake passage, and the intake passage inner wall surface portion around the throttle valve bulges outward and expands outward. The shape of the exposed inner wall surface portion of the intake passage is such that the flow passage area between the valve body peripheral edge of the throttle valve and the inner wall surface portion of the intake passage gradually increases as the throttle valve opens from the fully closed position. The internal combustion engine according to claim 1. 15. The internal combustion engine according to claim 1, wherein the exhaust gas recirculation rate is approximately 55% or more. 16. The internal combustion engine according to claim 1, wherein a catalyst having an oxidizing function is arranged in the engine exhaust passage. 17. The catalyst is an oxidation catalyst, a three-way catalyst or NO.
An internal combustion engine according to claim 16, wherein the _x absorbent comprises at least one. 18. The first combustion in which the amount of recirculated exhaust gas supplied to the combustion chamber is larger than the amount of recirculated exhaust gas at which the soot generation amount reaches a peak, and the soot generation amount peaks. 2. The internal combustion engine according to claim 1, further comprising switching means for selectively switching between the second combustion in which the amount of recirculated exhaust gas supplied to the combustion chamber is smaller than the amount of recirculated gas that becomes 19. The engine operating region is divided into a first operating region on the low load side and a second operating region on the high load side, and first combustion is performed in the first operating region to perform the second operating region. Second in the area
The internal combustion engine according to claim 18, wherein the combustion is performed.